This invention relates to two-phase pumped thermal control systems which include an evaporator for evaporating coolant to form coolant vapor and a condenser for condensing the vapor back to liquid, and further including a reservoir for coolant inventory management.
Future spacecraft are expected to use higher power in their operating systems than heretofore, and will consequently require substantial heat rejection capability. Compliance with the heat transfer requirements of future spacecraft requires thermal control techniques with the capabilities of heat acquisition, transport, and rejection of large heat loads from high heat density sources, and for transporting the heat over long distances under varying operational conditions. A desirable thermal control system is that which uses two-phase coolant flow. Two-phase means that the heat transfer fluid or coolant is evaporated to form a vapor at the source of heat and condenses back to liquid form at the location where heat rejection takes place.
It is advantageous in designing heat exchangers and other portions of spacecraft thermal control subsystems, to select not only those which will work and are practical in a zero gravity or microgravity environment, but which in addition can be tested in an Earth gravity environment and work in that environment in approximately the same manner as in a microgravity environment. Copending patent applications Ser. Nos. 111,338 filed Oct. 22, 1987 and 219,115 filed July 15, 1988, both in the name of Fredley, describe pumped systems in which cold liquid is piped from a condenser to "cold" plates, which cool the source of heat, and at which the liquid coolant absorbs heat and evaporates. The vapor returns to a heat rejection device such as flat radiating panels which cool the vapor by electromagnetic radiation to space. Such electromagnetic radiation is also known as thermal radiation. It is desirable in some cases to avoid the use of a mechanical pump in order to circulate the liquid. The capillary pump system is advantageous because it requires no moving parts, relying only upon a capillary evaporator, but it has the disadvantage that it generates a pressure potential or head of only about one-half pound per square inch (1/2psi), and thus requires a system in which each of the components has a very low pressure drop.
The temperature at which a coolant fluid makes a transition from its vapor state to a liquid state depends upon the pressure to which the coolant is subjected. Thus, the designer of such a system must select a system pressure which, together with other system operating parameters, gives the desired heat transfer and rejection performance. Consequently, the system pressure must be maintained within certain preselected limits. In general, the pressure within a closed system which includes fluid in both the vapor and liquid states depends upon the temperature of the closed loop.
FIG. 1a illustrates in simplified form a closed loop thermal control system 10. In FIG. 1a, a mechanical pump 12 pumps coolant liquid through a pipe 24 from a radiator or condenser 20 to an evaporator 14. Normally, the liquid leaving radiator 20 is subcooled or cooled to a temperature below the temperature at which vaporization takes place at the system pressure. This avoids the possibility of the coolant flashing to a vapor state the moment it enters the evaporator 14. Evaporator 14 is thermally coupled to a heat source 16 and transfers the heat from source 16 to the coolant as it traverses the evaporator. At some point within the evaporator 14, the coolant absorbs sufficient heat so that its sensible heat increases to the vaporization temperature, and absorption of additional heat causes the coolant to vaporize. Coolant vapor exits from the evaporator 14 by way of a vapor pipe 18, which carries the heat-laden coolant vapor to a radiator 20. The radiator 20 ordinarily includes large panels to which the heat from the vapor is transferred and which generates black-body radiation illustrated by a photon symbol 22. The black-body radiation transfers the heat to outer space, thereby cooling the vapor until, at some point along the radiator, it condenses into liquid form. The liquid coolant leaves the radiator 20 by way of pipe 24 for return to the evaporator.
FIG. 1b illustrates a heat transfer loop 30 which is generally similar to transfer loop 10 of FIG. 1a. Elements of FIGURE lb corresponding to those of FIG. 1a are designated by the same reference numerals. In FIG. 1b, however, there is no mechanical pump such as pump 12 of FIG. 1a. Instead, heat source 16 is coupled to a capillary evaporator 26 which generates a relatively small pressure head such as one-half pound per square inch. Such a system has reliability advantages because it lacks moving parts which might fail.
As mentioned in conjunction with FIG. 1a, the liquid coolant leaving the radiator 20 must be subcooled in order to prevent a situation in which liquid coolant flashes into vapor form as soon as it enters the evaporator 14 or 26. Such flashing may be disadvantageous, because it reduces the heat transfer capability of the evaporator. This is because a relatively large amount of heat is required to make a substantially constant-temperature transition between the liquid and vapor states, whereas it is difficult to transfer large amounts of heat to coolant vapor. Those skilled in the art know that the fluid pressure in a closed system such as those of FIG. 1 depends in part upon the environment and the load. For example, if radiator 20 is located in an environment in which it is "looking" or radiating into the cold environment of dark space, the system fluid pressure will tend to be lower than in the case in which it is in a warm environment. Also, an increase in the heat generated by the load, without other changes in the system environment, will tend to increase the system temperature and therefore the pressure.
FIG. 2a illustrates some details of a radiator 20 of FIG. 1 operating in a relatively cool environment. In FIG. 2a, vapor pipe 18 carries heat-laden coolant vapor to a manifold or plenum 36, which may be mounted on radiator 20. Manifold 36 distributes the vapor among a number, illustrated as five, pipes 38a, 38b, 38c, 38d and 38e, which are thermally coupled to radiator 20. A further manifold 40 collects condensed liquid from pipes 38a through 38e and transfers the liquid coolant to pipe 24. As illustrated in FIG. 2a, radiator 20 is in a relatively cold or dark environment, and radiates thermally to the environment as suggested by photon symbols 22. Since radiator 20 is relatively cold, heat is transferred away from the heat-laden vapor relatively quickly, and condensation takes place at a location transverse to pipes 38a through 38e illustrated by dot-dash line 42.
FIG. 2b is similar to FIG. 2a, and corresponding elements are designated by the same reference numerals. In FIG. 2b, however, the radiator 20 is in a relatively warm environment. Such a warm environment might be a location in which the sun, illustrated as 41, radiates onto the panel 20, as suggested by lines 44. Such radiation tends to raise the temperature of the radiator 20, and consequently the heat-laden vapor leaving manifold 36 must travel farther through pipes 38a through 38e before it condenses to liquid form. The location at which the vapor-to-liquid transition occurs is suggested by dot-dash line 46. By comparison of the conditions of FIG. 2b with that of FIG. 2a, it may be noted that placing the radiator panel 20 in a warm environment has resulted in a change in the ratio of liquid to vapor in pipes 38a through 38e. Portions of each pipe which in the case of FIG. 2a were filled with liquid are filled with vapor in the case of FIG. 2b. Put another way, a portion of the liquid in the pipes in the case of FIG. 2a has been turned to vapor in the situation of FIG. 2b. This tends to increase the fluid pressure within the closed system. The increase in fluid pressure may adversely affect the performance of the heat transfer loop, and may tend to rupture the components of the loop.
The problem of changes in pressure due to environmental conditions can be solved as illustrated in FIG. 3a. FIG. 3a is similar to FIG. 1b, and corresponding elements are designated by the same reference numerals. In FIG. 3a, a coolant reservoir illustrated as 50 is connected to liquid pipe 24 via another pipe 52. The coolant reservoir includes coolant in both liquid and vapor forms. The coolant reservoir accepts liquid coolant from the closed thermal control loop when system pressure tends to increase due to a hot environment of radiator 20 or to an increased heat load, and supplies liquid coolant to the thermal control loop when the pressure tends to drop due to a relatively cool environment surrounding the radiator 20 or a decreased heat load. As described below, this transfer of fluid is performed by controlling the temperature of the coolant in the reservoir. For example, the reservoir may be heated in order to raise the temperature of the vapor coolant therein, which creates a small pressure differential which forces liquid out of the reservoir into the heat transfer loop. Similarly, by turning off the heaters and allowing black-body radiation to cool the reservoir and coolant within, the pressure differential is reversed, thereby allowing liquid coolant from the heat transfer loop to be drawn back into the reservoir.
Generally speaking, coolant reservoir 50 must accommodate a volume of coolant sufficient to maintain substantially constant pressure within the closed thermal control loop 10; 30 under all environmental conditions, be capable of being emptied and refilled many times during a flight mission, control the flow of fluid by electrical heating and radiation cooling, prevent the generation of vapor pockets which might adversely affect operation, and prevent vapor bubbles from entering the outlet pipe. Such vapor bubble might deprime evaporator 26. Furthermore, the reservoir must function under all ordinary spacecraft environments and be capable of testing in a Earth gravity environment. While prevention of vapor pockets is desired, it should be understood that a vapor bubble will ordinarily exist within the reservoir.
FIG. 3b illustrates the physical configuration of one embodiment of a coolant reservoir 50 depicted in a FIG. 3a. As illustrated in FIGURE 3b, reservoir 50 includes a closed cylindrical wall or container 53, with an electrical heater illustrated as a circumferential band 54 located at one end of the cylindrical container. A pipe 52 connects the interior of container 53 with liquid pipe 24 of FIG. 3a. Wires illustrated as 56 connect heater 54 to a controlled source of power (not illustrated). FIG. 3c is an elevation view, partially cut away to reveal interior details, of reservoir 53 of FIG. 3b. Reservoir 53 is centered on an axis 8. FIG. 3d is a cross-section of the structure of FIG. 3c taken along section line d--d, orthogonal to axis 8.
Referring now to FIGS. 3c and 3d, a flat wick 60 extends transversely across the interior of container 53 between the cylindrical portion and curved end wall 78 of container 53 to thereby define a chamber 76. A pipe extension 152 continues pipe 52 and carries it to end within wick 60. Within cylindrical container 53, a metallic bracket illustrated as 66 joins the interior surface of container 53 along junction lines illustrated as 68 and 70. Bracket 66 together with the associated portion of container 53 defines a chamber 80 (on the left side in FIG. 3d) which is closed off at one end by flat wick 60. Bracket 66 is bent along two longitudinal lines 72 and 74 to define a flat central portion 82. An elongated member 84 is affixed to flat portion 82 of bracket 66 within chamber 80, and is shaped to define a longitudinal channel 85 which also terminates at wick 60. A further bracket 86 similar to bracket 66 is fastened within container 53. Bracket 86 also includes two bends to define a flat region 92 which is located opposite to flat region 82. Bracket 86 defines a further chamber 90 which is closed off or ends at wick 60. Chamber 90 is on the right side in FIGURE 3d. A further elongated member 94 affixed to the flat portion of bracket 86 defines a further longitudinal channel 95 which is closed off by wick 60. The interior wall of container 53 is lined with a circumferential wick, divided into a portion 62 in the upper half and a portion 64 in the lower half as illustrated in FIGS. 3c and 3d, and further portions 102 and 104 within chambers 80 and 90, respectively. Circumferential wick 102 within left side chamber 80 is turned near edges 68 and 70 and laps onto bracket 66, for reasons described in more detail below, at locations 113 and 114. A similar overlap occurs at 116 and 118 for wick 104 in right side chamber 90. A wick 106 covers that surface of bracket 66 which is remote from or faces away from chamber 80, and a similar wick 108 covers that surface of bracket 86 remote from chamber 90. For clarity in showing certain details described below, wick 106 is not shown in FIG. 3c. The surfaces of wicks 62 and 64 are represented by large cross-hatching. The facing flat portions 82 and 92 of brackets 66 and 86, respectively, may be near enough to each other so that wicks 106 and 108 are touching near axis 8. Brackets 66 and 86 together with the wall of container 53 define an upper chamber 81 and a lower chamber 83.
As so far generally described, liquid coolant can transfer between pipe 52 and interior chambers 76, 80, 81, 83 and 90 of reservoir 50, in both filling and emptying modes. To avoid redundancy, only an emptying mode is described in detail. Liquid coolant within chamber 76 is in contact with wick 60, and can leave chamber 76 by migration onto wick 60, thence to pipe extension 152 and through a wicked plug 112 to pipe 52. Wicked plug 112 aids in preventing the exit of any coolant vapor. Also, liquid can leave chamber 76 directly through an aperture 154 in pipe extension 152, which is occluded by wicked plug 112. Liquid in upper and lower chambers 81 and 83, and in side chambers 80 and 90, is in direct contact with wick 60 and can migrate therethrough to pipe extension 152 and thence to pipe 52. Liquid coolant in upper and lower chambers 81 and 83 is likely to be in contact with circumferential wicks 62 and 64, respectively, which in turn both communicate with wicks 106 and 108, which communicate with central longitudinal channels 85 and 95. Liquid within upper chamber 81 and lower chamber 83 can therefore migrate through wicks 62 and 64, and through wicks 106 and 108 to channels 85 and 95, and through the channels to end wick 60. From end wick 60, the liquid exits to pipe 52 as described above. The resistance to the flow of liquid is substantially greater within the wicks than in liquid channels such as 85 and 95. Consequently, upper chamber 81 tends to empty by lateral flow of liquid through its enveloping wicks 62, 106, and 108 to longitudinal channels 85 and 95, rather than by longitudinal flow through the enveloping wicks. Lower chamber 83 empties in the same fashion. As so far described, however, there is no communication between side chambers 80 and 90 and longitudinal channels 85 and 95. In the absence of such communication, side chambers 80 and 90 would have to empty by longitudinal flow through circumferential wick portions 102 and 104, which has relatively high resistance which might prevent fast emptying. Fast emptying in response to a small pressure differential is desirable for rapid response to changes in environmental conditions.
FIG. 3e is a cross-section of the structure of FIG. 3c taken along section lines e--e. Elements of FIG. 3e corresponding to those of FIGS. 3c and 3d are designated by the same reference numerals. Section line e--e of FIG. 3c is centered on a slot 122e formed in bracket 66. Slot 122e is one of a plurality of slots 122a, b, c, d, e, f . . . n, not all of which are visible in FIG. 3c. As mentioned above, wick 106 is not illustrated in FIG. 3C, so that slots 122a through 122f may be seen. As illustrated in FIG. 3c, portions of elongated member 84 can be seen through slots 122a through 122f, as can portions of longitudinal channel 85. Although not illustrated in FIG. 3c, circumferential wick 102 should be visible through the upper and lower portions of slots 122a through 122f. Each of slots 122a through 122f is dimensioned so that the surface tension forces of the coolant liquid in contact with the edges of the slot allow the coolant liquid to bridge across the slot, thereby forming a relatively low-resistance channel for the lateral flow of liquid. The channels defined by slots 122a through 122n bridge across longitudinal channels 85 and 95, thereby providing low-resistance channels for the flow of liquid which extend from the outer edges of the slots 122 to end wick 60, which closes off longitudinal channels 85 and 95. Consequently, upper and lower chambers 81 and 83, respectively, can transfer liquid at a relatively high rate to output wick 60, even at relatively low pressure differentials. In addition, since slots 122a through 122n pierce through bracket 66, they further provide lateral communication between side chambers 80 and 90 and the remainder of the system. In particular, bulk liquid can directly enter the low resistance channel defined by slot 122e from side chamber 80. In addition, the overlap portions 113 and 114 of wick 102 extend over the end of the slots, and in particular over the end of slot 122e, as illustrated in FIG. 3e, thereby providing communication between circumferential wick 102 and a low-resistance channel to the end wick 60.
As mentioned, it is very desirable that a thermal control system for operation in microgravity be capable of being tested in a gravity environment such as the Earth. Unlike the situation in the microgravity environment of space, the position or location of the liquid coolant is known to be at the bottom of the reservoir in a gravity environment. Depending upon the coolant, wicks having various mesh screens can raise the coolant liquid by different distances. For example, a 200-mesh aluminum screen can raise liquid ammonia (NH.sub.3) by about 5 inches. The finest mesh screen which is conveniently usable is 400-mesh, which can raise the liquid somewhat higher. Since the greatest heat transfer occurs as a result of the transition between liquid and vapor states, testing for maximum heat transfer capacity requires that liquid be available at the heat transfer locations. Thus, the wick must be wetted by liquid at all heat transfer locations. In space, a large vapor bubble might be located at either the right or left ends as viewed in FIG. 3c. This could be simulated in an Earth-gravity environment by orienting the reservoir with exit pipe 52 at the top or at the bottom, respectively. However, it may not be possible to perform operating tests in an Earth gravity environment for all ullage or fill conditions and at all orientations of the reservoir. This is because the testable ullage condition is limited to that at which the wick can raise liquid coolant to the highest point of heat transfer, which for liquid ammonia and 200-mesh screen is about 5 inches. When the liquid drops more than 5 inches below the heat transfer region, at least a portion of the wick in the heat transfer region goes dry. In a dry-wick region, application of heat to the outer surface (or radiation of heat therefrom) of the surface no longer results in heat transfer of the same nature as that which would occur in a space environment. Consequently, tests of a reservoir such as that of FIG. 3 can be made at all conditions of the vapor space only when the cylinder is lying on its side, and then only if its diameter is less than the wicking height. In other orientations, i.e., pipe 52 up or down, the vapor space must have a height of less than the wicking height, and testing at other conditions cannot be relied upon. This testing problem is made more serious because the location of the vapor space may change shape and move about under the influence of acceleration of the spacecraft in response to orders to perform maneuvers.
An improved coolant reservoir is desired.